Method and apparatus to control clutch fill pressure in an electro-mechanical transmission

ABSTRACT

A method to control a powertrain including a transmission, an engine, and an electric machine includes applying through a series of clutch fill events a series of incrementally changing command pressures in a pressure control solenoid controllably connected to a clutch within the transmission, monitoring a pressure switch fluidly connected to the pressure control solenoid and configured to indicate when the pressure switch is in a full feed state, determining changes in cycle times of the pressure switch corresponding to sequential applications of the series of incrementally changing command pressures, selecting a preferred command pressure to achieve a transient state in the clutch based upon the changes in pressure switch cycle times, and controlling the clutch based upon the preferred command pressure.

CROSS REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of U.S. Provisional Application No.60/982,874 filed on Oct. 26, 2007 which is hereby incorporated herein byreference.

TECHNICAL FIELD

This disclosure pertains to control systems for electro-mechanicaltransmissions.

BACKGROUND

The statements in this section merely provide background informationrelated to the present disclosure and may not constitute prior art.

Known powertrain architectures include torque-generative devices,including internal combustion engines and electric machines, whichtransmit torque through a transmission device to an output member. Oneexemplary powertrain includes a two-mode, compound-split,electro-mechanical transmission which utilizes an input member forreceiving motive torque from a prime mover power source, preferably aninternal combustion engine, and an output member. The output member canbe operatively connected to a driveline for a motor vehicle fortransmitting tractive torque thereto. Electric machines, operative asmotors or generators, generate a torque input to the transmission,independently of a torque input from the internal combustion engine. Theelectric machines may transform vehicle kinetic energy, transmittedthrough the vehicle driveline, to electrical energy that is storable inan electrical energy storage device. A control system monitors variousinputs from the vehicle and the operator and provides operationalcontrol of the powertrain, including controlling transmission operatingstate and gear shifting, controlling the torque-generative devices, andregulating the electrical power interchange among the electrical energystorage device and the electric machines to manage outputs of thetransmission, including torque and rotational speed. A hydraulic controlsystem is known to provide pressurized hydraulic oil for a number offunctions throughout the powertrain.

Operation of the above devices within a hybrid powertrain vehiclerequire management of numerous torque bearing shafts or devicesrepresenting connections to the above mentioned engine, electricalmachines, and driveline. Input torque from the engine and input torquefrom the electric machine or electric machines can be appliedindividually or cooperatively to provide output torque. Various controlschemes and operational connections between the various aforementionedcomponents of the hybrid drive system are known, and the control systemmust be able to engage to and disengage the various components from thetransmission in order to perform the functions of the hybrid powertrainsystem. Engagement and disengagement are known to be accomplished withinthe transmission by employing selectively operable clutches.

Clutches are devices well known in the art for engaging and disengagingshafts including the management of rotational velocity and torquedifferences between the shafts. Clutches are known in a variety ofdesigns and control methods. One known type of clutch is a mechanicalclutch operating by separating or joining two connective surfaces, forinstance, clutch plates, operating, when joined, to apply frictionaltorque to each other. One control method for operating such a mechanicalclutch includes applying the hydraulic control system implementingfluidic pressures transmitted through hydraulic lines to exert orrelease clamping force between the two connective surfaces. Operatedthusly, the clutch is not operated in a binary manner, but rather iscapable of a range of engagement states, from fully disengaged, tosynchronized but not engaged, to engaged but with only minimal clampingforce, to engaged with some maximum clamping force. The clamping forceavailable to be applied to the clutch determines how much reactivetorque the clutch can transmit before the clutch slips.

The hydraulic control system, as described above, utilizes lines chargedwith hydraulic oil to selectively activate clutches within thetransmission. Hydraulic switches or pressure control solenoids (PCS) areused to selectively apply pressure within a hydraulic control system. APCS can be electrically controlled, for instance with a magneticallyactuated solenoid device, well known in the art. Alternatively, a PCScan be hydraulically controlled, for example, actuated by a commandpressure and a return spring. Features within the PCS selectivelychannel or block hydraulic oil from passing therethrough depending uponthe actuation state of the PCS. In a blocked state, a PCS is known toinclude an exhaust path, allowing any trapped hydraulic oil to escape,thereby de-energizing the connected hydraulic circuit in order tocomplete the actuation cycle. Modulation of the command pressure canenable the PCS to be linearly, variably actuated, including actuationcontrolling application of fill pressure to the clutch in order toachieve within the clutch some middle or transient state between fullfeed and exhaust states.

A hydraulically actuated clutch operates by receiving pressurizedhydraulic oil into a clutch volume chamber. Hydraulic oil in this clutchvolume chamber exerts pressure upon features within the volume chamber.A piston or similar structure is known to be utilized to transform thishydraulic pressure into an articulation, for example a translatingmotion or compressing force. In an exemplary hydraulically actuatedclutch, pressurized hydraulic oil is used to fill a clutch volumechamber and thereby displace a clutch piston in order to selectivelyapply a compression force to the connective surfaces of the clutch. Arestoring force, for example as provided by a return spring, is known tobe used to counter the compressive force of the hydraulic oil. Asdescribed above, clutches are known to be engaged through a range ofengagement states. An exemplary clutch with all hydraulic pressureremoved can be in an unlocked state. An exemplary clutch with maximumhydraulic pressure can be in a locked state. An exemplary clutch withsome partial hydraulic force applied, wherein the force of the hydraulicoil and the force of a return spring are substantially equal, the clutchcan be in a touching state, with the plates in contact but with littleor no clamping force applied.

An engagement of a clutch, accomplished through a clutch fill event, isknown to be accomplished as rapidly as possible, with some minimumhydraulic pressure being maintained to assure rapid flow of thehydraulic oil into the clutch volume. However, rapid engagement of aclutch can cause a perceptible bump in the vehicle and cause shortenedlife of the component involved. A shock absorbing device can be utilizedto dampen the force of the rapid fill of the clutch volume chamber uponthe clutch. For example, a wave plate including a spring feature can beused between the cylinder piston and the clutch to absorb rapidincreases in hydraulic pressure. The touching state described above canbe defined as the clutch filled with enough hydraulic oil to cause zeroforce contact of the wave plate.

Clutch actuation status, such as would be indicated by a position sensoron a piston, is frequently not directly monitored. Such sensors wouldtend to be expensive, inaccurate, and increase warranty concerns.However, as described above, clutch operation in synchronous operationthrough a plurality of actuation states is a complicated processincluding multiple overlapping steps and control strategies. A methodallowing utilization of a command pressure to precisely actuate a clutchto an important transient state such as a touching state would bebeneficial.

SUMMARY

A powertrain includes an electro-mechanical transmissionmechanically-operatively coupled to an internal combustion engine and anelectric machine adapted to selectively transmit mechanical power to anoutput member via selective application of a plurality ofhydraulically-applied torque transfer clutches. A method to control thepowertrain includes applying through a series of clutch fill events aseries of incrementally changing command pressures in a pressure controlsolenoid controllably connected to one of the clutches, monitoring apressure switch fluidly connected to the pressure control solenoid andconfigured to indicate when the pressure switch is in a full feed state,determining changes in cycle times of the pressure switch correspondingto sequential applications of the series of incrementally changingcommand pressures, selecting a preferred command pressure to achieve atransient state in the clutch based upon the changes in pressure switchcycle times, and controlling the clutch based upon the preferred commandpressure.

BRIEF DESCRIPTION OF THE DRAWINGS

One or more embodiments will now be described, by way of example, withreference to the accompanying drawings, in which:

FIG. 1 is a schematic diagram of an exemplary powertrain, in accordancewith the present disclosure;

FIG. 2 is a schematic diagram of an exemplary architecture for a controlsystem and powertrain, in accordance with the present disclosure;

FIG. 3 is a schematic diagram of a hydraulic circuit, in accordance withthe present disclosure;

FIG. 4 schematically illustrates an exemplary clutch control circuitutilizing a hydraulically activated pressure control switch, inaccordance with the present disclosure;

FIG. 5 schematically illustrates an exemplary hydraulically actuatedclutch operated to provide clamping force upon a mechanical clutch, inaccordance with the present disclosure;

FIG. 6 graphically illustrates fill times required to reach an overlapstate versus the command pressure in the PCS, in accordance with thepresent disclosure;

FIG. 7 graphically illustrates fill times required to reach an overlapstate versus the command pressure through a range of incrementallydecreased command pressures, in accordance with the present disclosure;and

FIG. 8 illustrates a flowchart describing an exemplary process to selectand continually validate preferred command pressures, in accordance withthe disclosure.

DETAILED DESCRIPTION

Referring now to the drawings, wherein the showings are for the purposeof illustrating certain exemplary embodiments only and not for thepurpose of limiting the same, FIGS. 1 and 2 depict an exemplaryelectro-mechanical hybrid powertrain. The exemplary electro-mechanicalhybrid powertrain in accordance with the present disclosure is depictedin FIG. 1, comprising a two-mode, compound-split, electro-mechanicalhybrid transmission 10 operatively connected to an engine 14 and firstand second electric machines (‘MG-A’) 56 and (‘MG-B’) 72. The engine 14and first and second electric machines 56 and 72 each generate powerwhich can be transmitted to the transmission 10. The power generated bythe engine 14 and the first and second electric machines 56 and 72 andtransmitted to the transmission 10 is described in terms of inputtorques, referred to herein as T_(I), T_(A), and T_(B) respectively, andspeed, referred to herein as N_(I), N_(A), and N_(B), respectively.

The exemplary engine 14 comprises a multi-cylinder internal combustionengine selectively operative in several states to transmit torque to thetransmission 10 via an input shaft 12, and can be either aspark-ignition or a compression-ignition engine. The engine 14 includesa crankshaft (not shown) operatively coupled to the input shaft 12 ofthe transmission 10. A rotational speed sensor 11 monitors rotationalspeed of the input shaft 12. Power output from the engine 14, comprisingrotational speed and output torque, can differ from the input speed,N_(I), and the input torque, T_(I), to the transmission 10 due toplacement of torque-consuming components on the input shaft 12 betweenthe engine 14 and the transmission 10, e.g., a hydraulic pump (notshown) and/or a torque management device (not shown).

The exemplary transmission 10 comprises three planetary-gear sets 24, 26and 28, and four selectively engageable torque-transmitting devices,i.e., clutches C1 70, C2 62, C3 73, and C4 75. As used herein, clutchesrefer to any type of friction torque transfer device including single orcompound plate clutches or packs, band clutches, and brakes, forexample. A hydraulic control circuit 42, preferably controlled by atransmission control module (hereafter ‘TCM’) 17, is operative tocontrol clutch states. Clutches C2 62 and C4 75 preferably comprisehydraulically-applied rotating friction clutches. Clutches C1 70 and C373 preferably comprise hydraulically-controlled stationary devices thatcan be selectively grounded to a transmission case 68. Each of theclutches C1 70, C2 62, C3 73, and C4 75 is preferably hydraulicallyapplied, selectively receiving pressurized hydraulic oil via thehydraulic control circuit 42.

The first and second electric machines 56 and 72 preferably comprisethree-phase AC machines, each including a stator (not shown) and a rotor(not shown), and respective resolvers 80 and 82. The motor stator foreach machine is grounded to an outer portion of the transmission case68, and includes a stator core with coiled electrical windings extendingtherefrom. The rotor for the first electric machine 56 is supported on ahub plate gear that is operatively attached to shaft 60 via the secondplanetary gear set 26. The rotor for the second electric machine 72 isfixedly attached to a sleeve shaft hub 66.

Each of the resolvers 80 and 82 preferably comprises a variablereluctance device including a resolver stator (not shown) and a resolverrotor (not shown). The resolvers 80 and 82 are appropriately positionedand assembled on respective ones of the first and second electricmachines 56 and 72. Stators of respective ones of the resolvers 80 and82 are operatively connected to one of the stators for the first andsecond electric machines 56 and 72. The resolver rotors are operativelyconnected to the rotor for the corresponding first and second electricmachines 56 and 72. Each of the resolvers 80 and 82 is signally andoperatively connected to a transmission power inverter control module(hereafter ‘TPIM’) 19, and each senses and monitors rotational positionof the resolver rotor relative to the resolver stator, thus monitoringrotational position of respective ones of first and second electricmachines 56 and 72. Additionally, the signals output from the resolvers80 and 82 are interpreted to provide the rotational speeds for first andsecond electric machines 56 and 72, i.e., N_(A) and N_(B), respectively.

The transmission 10 includes an output member 64, e.g. a shaft, which isoperably connected to a driveline 90 for a vehicle (not shown), toprovide output power, e.g., to vehicle wheels 93, one of which is shownin FIG. 1. The output power is characterized in terms of an outputrotational speed, N_(O) and an output torque, T_(O). A transmissionoutput speed sensor 84 monitors rotational speed and rotationaldirection of the output member 64. Each of the vehicle wheels 93, ispreferably equipped with a sensor 94 adapted to monitor wheel speed,V_(SS-WHL), the output of which is monitored by a control module of adistributed control module system described with respect to FIG. 2, todetermine vehicle speed, and absolute and relative wheel speeds forbraking control, traction control, and vehicle acceleration management.

The input torques from the engine 14 and the first and second electricmachines 56 and 72 (T_(I), T_(A), and T_(B) respectively) are generatedas a result of energy conversion from fuel or electrical potentialstored in an electrical energy storage device (hereafter ‘ESD’) 74. TheESD 74 is high voltage DC-coupled to the TPIM 19 via DC transferconductors 27. The transfer conductors 27 include a contactor switch 38.When the contactor switch 38 is closed, under normal operation, electriccurrent can flow between the ESD 74 and the TPIM 19. When the contactorswitch 38 is opened electric current flow between the ESD 74 and theTPIM 19 is interrupted. The TPIM 19 transmits electrical power to andfrom the first electric machine 56 by transfer conductors 29, and theTPIM 19 similarly transmits electrical power to and from the secondelectric machine 72 by transfer conductors 31, in response to torquecommands for the first and second electric machines 56 and 72 to achievethe input torques T_(A) and T_(B). Electrical current is transmitted toand from the ESD 74 in accordance with whether the ESD 74 is beingcharged or discharged.

The TPIM 19 includes the pair of power inverters (not shown) andrespective motor control modules (not shown) configured to receive thetorque commands and control inverter states therefrom for providingmotor drive or regeneration functionality to meet the commanded motortorques T_(A) and T_(B). The power inverters comprise knowncomplementary three-phase power electronics devices, and each includes aplurality of insulated gate bipolar transistors (not shown) forconverting DC power from the ESD 74 to AC power for powering respectiveones of the first and second electric machines 56 and 72, by switchingat high frequencies. The insulated gate bipolar transistors form aswitch mode power supply configured to receive control commands. Thereis typically one pair of insulated gate bipolar transistors for eachphase of each of the three-phase electric machines. States of theinsulated gate bipolar transistors are controlled to provide motor drivemechanical power generation or electric power regenerationfunctionality. The three-phase inverters receive or supply DC electricpower via DC transfer conductors 27 and transform it to or fromthree-phase AC power, which is conducted to or from the first and secondelectric machines 56 and 72 for operation as motors or generators viatransfer conductors 29 and 31 respectively.

FIG. 2 is a schematic block diagram of the distributed control modulesystem. The elements described hereinafter comprise a subset of anoverall vehicle control architecture, and provide coordinated systemcontrol of the exemplary powertrain described in FIG. 1. The distributedcontrol module system synthesizes pertinent information and inputs, andexecutes algorithms to control various actuators to achieve controlobjectives, including objectives related to fuel economy, emissions,performance, drivability, and protection of hardware, includingbatteries of ESD 74 and the first and second electric machines 56 and72. The distributed control module system includes an engine controlmodule (hereafter ‘ECM’) 23, the TCM 17, a battery pack control module(hereafter ‘BPCM’) 21, and the TPIM 19. A hybrid control module(hereafter ‘HCP’) 5 provides supervisory control and coordination of theECM 23, the TCM 17, the BPCM 21, and the TPIM 19. A user interface(‘UI’) 13 is operatively connected to a plurality of devices throughwhich a vehicle operator controls or directs operation of theelectro-mechanical hybrid powertrain. The devices include an acceleratorpedal 113 (‘AP’) from which an operator torque request is determined, anoperator brake pedal 112 (‘BP’), a transmission gear selector 114(‘PRNDL’), and a vehicle speed cruise control (not shown). Thetransmission gear selector 114 may have a discrete number ofoperator-selectable positions, including the rotational direction of theoutput member 64 to enable one of a forward and a reverse direction.

The aforementioned control modules communicate with other controlmodules, sensors, and actuators via a local area network (hereafter‘LAN’) bus 6. The LAN bus 6 allows for structured communication ofstates of operating parameters and actuator command signals between thevarious control modules. The specific communication protocol utilized isapplication-specific. The LAN bus 6 and appropriate protocols providefor robust messaging and multi-control module interfacing between theaforementioned control modules, and other control modules providingfunctionality such as antilock braking, traction control, and vehiclestability. Multiple communications buses may be used to improvecommunications speed and provide some level of signal redundancy andintegrity. Communication between individual control modules can also beeffected using a direct link, e.g., a serial peripheral interface(‘SPI’) bus (not shown).

The HCP 5 provides supervisory control of the powertrain, serving tocoordinate operation of the ECM 23, TCM 17, TPIM 19, and BPCM 21. Basedupon various input signals from the user interface 13 and thepowertrain, including the ESD 74, the HCP 5 generates various commands,including: the operator torque request (‘T_(O) _(—) _(REQ)’), acommanded output torque (‘T_(CMD)’) to the driveline 90, an engine inputtorque command, clutch torques for the torque-transfer clutches C1 70,C2 62, C3 73, C4 75 of the transmission 10; and the torque commands forthe first and second electric machines 56 and 72, respectively. The TCM17 is operatively connected to the hydraulic control circuit 42 andprovides various functions including monitoring various pressure sensingdevices (not shown) and generating and communicating control signals tovarious solenoids (not shown) thereby controlling pressure switches andcontrol valves contained within the hydraulic control circuit 42.

The ECM 23 is operatively connected to the engine 14, and functions toacquire data from sensors and control actuators of the engine 14 over aplurality of discrete lines, shown for simplicity as an aggregatebi-directional interface cable 35. The ECM 23 receives the engine inputtorque command from the HCP 5. The ECM 23 determines the actual engineinput torque, T_(I), provided to the transmission 10 at that point intime based upon monitored engine speed and load, which is communicatedto the HCP 5. The ECM 23 monitors input from the rotational speed sensor11 to determine the engine input speed to the input shaft 12, whichtranslates to the transmission input speed, N_(I). The ECM 23 monitorsinputs from sensors (not shown) to determine states of other engineoperating parameters including, e.g., a manifold pressure, enginecoolant temperature, ambient air temperature, and ambient pressure. Theengine load can be determined, for example, from the manifold pressure,or alternatively, from monitoring operator input to the acceleratorpedal 113. The ECM 23 generates and communicates command signals tocontrol engine actuators, including, e.g., fuel injectors, ignitionmodules, and throttle control modules, none of which are shown.

The TCM 17 is operatively connected to the transmission 10 and monitorsinputs from sensors (not shown) to determine states of transmissionoperating parameters. The TCM 17 generates and communicates commandsignals to control the transmission 10, including controlling thehydraulic control circuit 42. Inputs from the TCM 17 to the HCP 5include estimated clutch torques for each of the clutches, i.e., C1 70,C2 62, C3 73, and C4 75, and rotational output speed, N_(O), of theoutput member 64. Other actuators and sensors may be used to provideadditional information from the TCM 17 to the HCP 5 for controlpurposes. The TCM 17 monitors inputs from pressure switches (not shown)and selectively actuates pressure control solenoids (not shown) andshift solenoids (not shown) of the hydraulic control circuit 42 toselectively actuate the various clutches C1 70, C2 62, C3 73, and C4 75to achieve various transmission operating range states, as describedhereinbelow.

The BPCM 21 is signally connected to sensors (not shown) to monitor theESD 74, including states of electrical current and voltage parameters,to provide information indicative of parametric states of the batteriesof the ESD 74 to the HCP 5. The parametric states of the batteriespreferably include battery state-of-charge, battery voltage, batterytemperature, and available battery power, referred to as a rangeP_(BAT)_MIN to P_(BAT)_MAX.

Each of the control modules ECM 23, TCM 17, TPIM 19 and BPCM 21 ispreferably a general-purpose digital computer comprising amicroprocessor or central processing unit, storage mediums comprisingread only memory (‘ROM’), random access memory (‘RAM’), electricallyprogrammable read only memory (‘EPROM’), a high speed clock, analog todigital (‘A/D’) and digital to analog (‘D/A’) circuitry, andinput/output circuitry and devices (‘I/O’) and appropriate signalconditioning and buffer circuitry. Each of the control modules has a setof control algorithms, comprising resident program instructions andcalibrations stored in one of the storage mediums and executed toprovide the respective functions of each computer. Information transferbetween the control modules is preferably accomplished using the LAN bus6 and SPI buses. The control algorithms are executed during preset loopcycles such that each algorithm is executed at least once each loopcycle. Algorithms stored in the non-volatile memory devices are executedby one of the central processing units to monitor inputs from thesensing devices and execute control and diagnostic routines to controloperation of the actuators, using preset calibrations. Loop cycles areexecuted at regular intervals, for example each 3.125, 6.25, 12.5, 25and 100 milliseconds during ongoing operation of the powertrain.Alternatively, algorithms may be executed in response to the occurrenceof an event.

The exemplary powertrain selectively operates in one of severaloperating range states that can be described in terms of an engine statecomprising one of an engine on state (‘ON’) and an engine off state(‘OFF’), and a transmission state comprising a plurality of fixed gearsand continuously variable operating modes, described with reference toTable 1, below.

TABLE 1 Engine Transmission Operating Applied Description State RangeState Clutches MI_Eng_Off OFF EVT Mode I C1 70 MI_Eng_On ON EVT Mode IC1 70 FG1 ON Fixed Gear Ratio 1 C1 70 C4 75 FG2 ON Fixed Gear Ratio 2 C170 C2 62 MII_Eng_Off OFF EVT Mode II C2 62 MII_Eng_On ON EVT Mode II C262 FG3 ON Fixed Gear Ratio 3 C2 62 C4 75 FG4 ON Fixed Gear Ratio 4 C2 62C3 73

Each of the transmission operating range states is described in thetable and indicates which of the specific clutches C1 70, C2 62, C3 73,and C4 75 are applied for each of the operating range states. A firstcontinuously variable mode, i.e., EVT Mode I, or MI, is selected byapplying clutch C1 70 only in order to “ground” the outer gear member ofthe third planetary gear set 28. The engine state can be one of ON(‘MI_Eng_On’) or OFF (‘MI_Eng_Off’). A second continuously variablemode, i.e., EVT Mode II, or MII, is selected by applying clutch C2 62only to connect the shaft 60 to the carrier of the third planetary gearset 28. The engine state can be one of ON (‘MII_Eng_On’) or OFF(‘MII_Eng_Off’). For purposes of this description, when the engine stateis OFF, the engine input speed is equal to zero revolutions per minute(‘RPM’), i.e., the engine crankshaft is not rotating. A fixed gearoperation provides a fixed ratio operation of input-to-output speed ofthe transmission 10, i.e., N_(I)/N_(O), is achieved. A first fixed gearoperation (‘FG1’) is selected by applying clutches C1 70 and C4 75. Asecond fixed gear operation (‘FG2’) is selected by applying clutches C170 and C2 62. A third fixed gear operation (‘FG3’) is selected byapplying clutches C2 62 and C4 75. A fourth fixed gear operation (‘FG4’)is selected by applying clutches C2 62 and C3 73. The fixed ratiooperation of input-to-output speed increases with increased fixed gearoperation due to decreased gear ratios in the planetary gears 24, 26,and 28. The rotational speeds of the first and second electric machines56 and 72, N_(A) and N_(B) respectively, are dependent on internalrotation of the mechanism as defined by the clutching and areproportional to the input speed measured at the input shaft 12.

In response to operator input via the accelerator pedal 113 and brakepedal 112 as captured by the user interface 13, the HCP 5 and one ormore of the other control modules determine the commanded output torque,T_(CMD), intended to meet the operator torque request, T_(O)_REQ, to beexecuted at the output member 64 and transmitted to the driveline 90.Final vehicle acceleration is affected by other factors including, e.g.,road load, road grade, and vehicle mass. The operating range state isdetermined for the transmission 10 based upon a variety of operatingcharacteristics of the powertrain. This includes the operator torquerequest, communicated through the accelerator pedal 113 and brake pedal112 to the user interface 13 as previously described. The operatingrange state may be predicated on a powertrain torque demand caused by acommand to operate the first and second electric machines 56 and 72 inan electrical energy generating mode or in a torque generating mode. Theoperating range state can be determined by an optimization algorithm orroutine which determines optimum system efficiency based upon operatordemand for power, battery state of charge, and energy efficiencies ofthe engine 14 and the first and second electric machines 56 and 72. Thecontrol system manages torque inputs from the engine 14 and the firstand second electric machines 56 and 72 based upon an outcome of theexecuted optimization routine, and system efficiencies are optimizedthereby, to manage fuel economy and battery charging. Furthermore,operation can be determined based upon a fault in a component or system.The HCP 5 monitors the torque-generative devices, and determines thepower output from the transmission 10 required to achieve the desiredoutput torque to meet the operator torque request. As should be apparentfrom the description above, the ESD 74 and the first and second electricmachines 56 and 72 are electrically-operatively coupled for power flowtherebetween. Furthermore, the engine 14, the first and second electricmachines 56 and 72, and the electro-mechanical transmission 10 aremechanically-operatively coupled to transmit power therebetween togenerate a power flow to the output member 64.

FIG. 3 depicts a schematic diagram of the hydraulic control circuit 42for controlling flow of hydraulic oil in the exemplary transmission. Amain hydraulic pump 88 is driven off the input shaft 12 from the engine14, and an auxiliary pump 110 controlled by the TPIM 19 to providepressurized fluid to the hydraulic control circuit 42 through valve 140.The auxiliary pump 110 preferably comprises an electrically-powered pumpof an appropriate size and capacity to provide sufficient flow ofpressurized hydraulic oil into the hydraulic control circuit 42 whenoperational. The hydraulic control circuit 42 selectively distributeshydraulic pressure to a plurality of devices, including thetorque-transfer clutches C1 70, C2 62, C3 73, and C4 75, active coolingcircuits for the first and second electric machines 56 and 72 (notshown), and a base cooling circuit for cooling and lubricating thetransmission 10 via passages 142, 144 (not depicted in detail). Aspreviously stated, the TCM 17 actuates the various clutches to achieveone of the transmission operating range states through selectiveactuation of hydraulic circuit flow control devices comprising variablepressure control solenoids (‘PCS’) PCS1 108, PCS2 114, PCS3 112, PCS4116 and solenoid-controlled flow management valves, X-valve 119 andY-valve 121. The hydraulic control circuit 42 is fluidly connected topressure switches PS1, PS2, PS3, and PS4 via passages 122, 124, 126, and128, respectively. The pressure control solenoid PCS1 108 has a controlposition of normally high and is operative to modulate the magnitude offluidic pressure in the hydraulic circuit through fluidic interactionwith controllable pressure regulator 107 and spool valve 109. Thecontrollable pressure regulator 107 and spool valve 109 interact withPCS1 108 to control hydraulic pressure in the hydraulic control circuit42 over a range of pressures and may provide additional functionalityfor the hydraulic control circuit 42. Pressure control solenoid PCS3 112has a control position of normally high, and is fluidly connected tospool valve 113 and operative to effect flow therethrough when actuated.Spool valve 113 is fluidly connected to pressure switch PS3 via passage126. Pressure control solenoid PCS2 114 has a control position ofnormally high, and is fluidly connected to spool valve 115 and operativeto effect flow therethrough when actuated. Spool valve 115 is fluidlyconnected to pressure switch PS2 via passage 124. Pressure controlsolenoid PCS4 116 has a control position of normally low, and is fluidlyconnected to spool valve 117 and operative to effect flow therethroughwhen actuated. Spool valve 117 is fluidly connected to pressure switchPS4 via passage 128.

The X-Valve 119 and Y-Valve 121 each comprise flow management valvescontrolled by solenoids 118, 120, respectively, in the exemplary system,and have control states of High (‘1’) and Low (‘0’). The control statesrefer to positions of each valve to which control flow to differentdevices in the hydraulic control circuit 42 and the transmission 10. TheX-valve 119 is operative to direct pressurized fluid to clutches C3 73and C4 75 and cooling systems for stators of the first and secondelectric machines 56 and 72 via fluidic passages 136, 138, 144, 142respectively, depending upon the source of the fluidic input, as isdescribed hereinafter. The Y-valve 121 is operative to directpressurized fluid to clutches C1 70 and C2 62 via fluidic passages 132and 134 respectively, depending upon the source of the fluidic input, asis described hereinafter. The Y-valve 121 is fluidly connected topressure switch PS1 via passage 122.

The hydraulic control circuit 42 includes a base cooling circuit forproviding hydraulic oil to cool the stators of the first and secondelectric machines 56 and 72. The base cooling circuit includes fluidconduits from the valve 140 flowing directly to a flow restrictor whichleads to fluidic passage 144 leading to the base cooling circuit for thestator of the first electric machine 56, and to a flow restrictor whichleads to fluidic passage 142 leading to the base cooling circuit for thestator of the second electric machine 72. Active cooling of stators forthe first and second electric machines 56 and 72 is effected byselective actuation of pressure control solenoids PCS2 114, PCS3 112 andPCS4 116 and solenoid-controlled flow management valves X-valve 119 andY-valve 121, which leads to flow of hydraulic oil around the selectedstator and permits heat to be transferred therebetween, primarilythrough conduction.

An exemplary logic table to accomplish control of the exemplaryhydraulic control circuit 42 to control operation of the transmission 10in one of the transmission operating range states is provided withreference to Table 2, below.

TABLE 2 X- Y- Transmission Valve Valve Operating Logic Logic PCS1 PCS2PCS3 PCS4 Range No C2 Normal Normal Normal Normal State Latch Latch HighHigh High Low EVT 0 0 Line MG-B C1 MG-A Mode I Modulation Stator StatorCool Cool EVT 0 1 Line C2 MG-B MG-A Mode II Modulation Stator StatorCool Cool Low 1 0 Line C2 C1 C4 Range Modulation High 1 1 Line C2 C3 C4Range Modulation

A Low Range is defined as a transmission operating range statecomprising one of the first continuously variable mode and the first andsecond fixed gear operations. A High Range is defined as a transmissionoperating range state comprising one of the second continuously variablemode and the third and fourth fixed gear operations. Selective controlof the X-valve 119 and the Y-valve 121 and actuation of the solenoidsPCS2 112, PCS3 114, PCS4 116 facilitate flow of hydraulic oil to actuateclutches C1 70, C2 63, C3 73, and C4 75, and provide cooling for thestators the first and second electric machines 56 and 72.

In operation, a transmission operating range state, i.e. one of thefixed gear and continuously variable mode operations, is selected forthe exemplary transmission 10 based upon a variety of operatingcharacteristics of the powertrain. This includes the operator torquerequest, typically communicated through inputs to the UI 13 aspreviously described. Additionally, a demand for output torque ispredicated on external conditions, including, e.g., road grade, roadsurface conditions, or wind load. The operating range state may bepredicated on a powertrain torque demand caused by a control modulecommand to operate of the electrical machines in an electrical energygenerating mode or in a torque generating mode. The operating rangestate can be determined by an optimization algorithm or routine operableto determine an optimum system efficiency based upon the operator torquerequest, battery state of charge, and energy efficiencies of the engine14 and the first and second electric machines 56 and 72. The controlsystem manages the input torques from the engine 14 and the first andsecond electric machines 56 and 72 based upon an outcome of the executedoptimization routine, and system optimization occurs to improve fueleconomy and manage battery charging. Furthermore, the operation can bedetermined based upon a fault in a component or system.

FIG. 4 schematically illustrates an exemplary clutch control circuitutilizing a hydraulically activated pressure control switch, inaccordance with the present disclosure. Clutch control circuit 200includes PCS 210, pressure switch 240, and hydraulic lines 270, 272,274, 276, 278, and 280. PCS 210 selectively controls flow of pressurizedhydraulic oil to and from a hydraulically actuated clutch (not shown) bytranslation of selecting mechanism within the PCS, in this exemplaryembodiment, a spool valve plunger 220. Plunger 220 is selectively actedupon from a first end 222 of the plunger and a second end 224 of theplunger, the balance of forces determining the translative position ofthe plunger within the PCS. Plunger 220 includes plunger details 226including holes, grooves, channels, or other features formed on theplunger in order to selectively direct hydraulic oil between variousports connecting hydraulic lines to PCS 210. The position of plunger 220within PCS 210, corresponding to clutch states described above,selectively align plunger details 226 with hydraulic lines accomplishingthe intended clutch function. In the exemplary clutch of FIG. 4, aplunger position to the right corresponds to a full feed state, whereinhydraulic pressure from a main pressure line 272 is channeled throughplunger details 226 to clutch feed line 276. Similarly, a plungerposition to the left corresponds to an exhaust state, wherein hydraulicoil within the clutch is allowed to escape the clutch and flow throughexhaust line 274, entering a hydraulic control system return line (notshown). Selecting the position of plunger 220 is accomplished bymodulating a command pressure to a command pressure line 270 feeding acommand pressure volume 260 in contact with first end 222. As will beappreciated by one having ordinary skill in the art, force created bypressure on a surface can be determined through the following equation:

FORCE=PRESSURE*SURFACE_AREA_ACTED_UPON   [1]

In the case of exemplary plunger 220, the force acting upon the plungerfrom the left equals the hydraulic pressure achieved within commandpressure volume 260 times the surface area of first end 222. An increasein pressure within command pressure volume 260 increases the forceacting upon plunger 220 from the side of first end 222. A valve returnspring 250 applies a force to the second end 224, acting as arestorative force in the opposite direction of the pressure withincommand pressure volume 260. Force resulting from pressure within volume260 and force from spring 250 act together such that increased pressurewithin command pressure volume 260 tends to move plunger 220 in onedirection, and reduced pressure within command pressure volume 260 tendsto move plunger 220 in the opposite direction. Exemplary PCS 210includes another feature including a feedback line 278. Hydraulic oilflowing through clutch feed line 276 additionally flows or applies apressure through feedback line 278. Hydraulic oil from feedback line 278re-enters PCS 210 within a feedback pressure volume 265 located on thesame side of plunger 220 as spring 250. Force resulting upon plunger 220from hydraulic pressure within feedback pressure volume 265 counteractsforce resulting from hydraulic pressure within command pressure volume260. As a result, wherein a balance of forces resulting from pressurewithin command pressure volume 260 and spring 250 would cause plunger220 to be in a position correlating to a full feed state, elevatedpressure achieved within clutch feed line 276 associated with a clutchfill event reaching a certain progression creates a force acting uponplunger 220 away from the full feed state position. Calibration and/orcontrol of feedback line 278 and resulting force upon plunger 220corresponding to a selected pressure within command pressure volume 260can create a self-correcting plunger position between the opposite endsof plunger travel, enabling an overlap state. Such an overlap state isuseful for modulating the pressure achieved within the clutch, forexample, enabling calibrated control to a touching state for the clutch.Full feed state can still be achieved despite operation of the feedbackline 278 by setting pressure within the command pressure volume 260 toapply a force to plunger 220 exceeding the combination of the forceapplied by spring 250 and force resulting from hydraulic pressure withinfeedback pressure volume 265. PCS 210 is known to include pressureswitch 240, fed by pressure switch line 280, utilized in known controlmethods to indicate pressure levels required for control of PCS 210. Inthis way, PCS 210 can selectively channel hydraulic oil to accomplishmultiple states within a hydraulically activated clutch.

By modulating a command pressure, a PCS of the above exemplaryconfiguration can operate in three states. A high command pressurecommands a full feed state, allowing full exposure of P_(LINE) to theclutch being filled. A low or null command pressure commands an exhauststate, blocking access of P_(LINE) to the clutch and providing a path toexhaust hydraulic pressure from within the clutch. An intermediate orcalibrated command pressure commands an overlap state. The function ofan overlap state depends upon the calibration of the calibrated commandpressure. An exemplary function of such an overlap state is to command atouching state in the clutch. Selective calibration of the commandpressure to achieve the overlap state, in combination with monitoredoperation of the pressure switch, allows for accurately selecting a filllevel within the clutch, for example, a fill level corresponding to atouching state in the clutch.

A number of PCS physical configurations are known. One exemplary PCSconfiguration, as described above, utilizes a cylindrical plungerlocated in a cylindrical housing. However, a multitude of shapes,configurations, activations methods, and calibration strategies areknown in the art, and this disclosure is not intended to be limited tothe particular exemplary embodiments described herein.

Pressure switch 240 is calibrated to indicate a pressure reaching somelevel. In the particular embodiment described in FIG. 4, the pressureswitch can be utilized for example, to indicate a positive signal onlywhen the PCS is in a full feed state. In such an exemplary use, thecalibration of the pressure switch indication need not correspond to theactual pressures to which it is exposed, for example pressure levels inthe command pressure volume 260, but can rather indicates some nominallevel which the pressure always exceeds when the pressure switch isexposed to the pressurized hydraulic fluid.

As described above, operation and control of clutches are important tooperating a complex powertrain, such as a hybrid powertrain.Drivability, fuel efficiency, and component life are all impacted by theoperation of clutches within the system. Known methods utilizing look-uptables to control clutch activating devices, such as a PCS, areimprecise and inefficient. Much can be determined within a hydrauliccontrol system based upon analysis of available inputs. A method isdisclosed for localizing a preferred command pressure to attain atouching state within a transmission clutch based upon clutch fill timesand pressure switch readings.

A hydraulically actuated clutch utilizes selectively actuatedpressurized hydraulic flow to create a desired motion or compression. Anexemplary clutch operates by receiving pressurized hydraulic oil into aclutch volume chamber. FIG. 5 schematically illustrates an exemplaryhydraulically actuated clutch operated to provide clamping force upon amechanical clutch, in accordance with the present disclosure. Clutchassembly 300 comprises a clutch cylinder 320 and a mechanical clutch340. Clutch cylinder 320 includes a piston 322 and a clutch volumechamber 324. Pressurized hydraulic fluid at some fill pressure entersclutch volume chamber 324 through hydraulic line 350. Hydraulic line 350is fluidly connected with a mechanism for selectively applying hydraulicflow, such as an exemplary PCS device (not shown). Hydraulic oil inclutch volume chamber 324 exerts pressure upon features within thevolume chamber. Piston 322 transforms the fill pressure exerted by thehydraulic fluid into a force. The force transmitted through piston 322is used to articulate mechanical clutch 340 through various statesrequired according to synchronous clutch operation described above.Positive hydraulic pressure is used to fill the clutch volume chamber324 and move piston 322 in one direction. As will be appreciated by onehaving ordinary skill in the art, evacuation of hydraulic oil fromclutch volume chamber 324 acts in some degree to move piston 322 in theother direction, but cavitation limits the ability of low pressurehydraulic fluid from effectively moving piston 322. As a result, returnspring 326 is utilized to provide force to move piston 322 in thedirection opposite to the direction achieved through the application ofpressurized hydraulic fluid.

Mechanical clutch 340 is selectively actuated by the transmission offorce through piston 322. Mechanical clutch 340 includes clutchconnective surfaces in the form of clutch plates 345. Clutch plates 345are connected to rotating members within the transmission (not shown).When mechanical clutch 340 is not actuated, clutch plates 345 are keptseparate. Spinning of some fraction of clutch plates 345 does not causespinning of the remaining fraction of clutch plates 345. When mechanicalclutch 340 is actuated, clutch plates 345 are brought into contact withneighboring plates, and sufficient frictional forces between clutchplates 345 creates a locked relationship wherein the plates move inunison. Between rotating objects applying a torque, the torque capacity(‘T_(C)’) generated between the objects can be determined by thefollowing equation:

$\begin{matrix}{T_{C} = {\frac{2}{3}*f*F_{A}}} & \lbrack 2\rbrack\end{matrix}$

f is the coefficient of friction between the rotating objects. As willbe appreciated by one having ordinary skill in the art, f changesdepending upon whether there is relative movement between the twoobjects. F_(A) is the axial force applied normally to direction ofrotation of the objects. F_(A) in mechanical clutch 340 is generated bycompressive force transmitted through piston 322. When the clutch is ina touching state, F_(A) is kept at substantially zero, yielding zerotorque capacity.

A process transitioning piston 322 from one extreme of motion to theother includes three distinct phases. A first phase, beginning from afully unlocked state in the clutch, wherein no hydraulic pressure isbeing applied upon piston 322, the exemplary piston 322 is in a fullyleft position, as depicted in FIG. 5, and has no contact with mechanicalclutch 340 or clutch plates 345. As pressurized hydraulic fluid at afill pressure is directed into clutch volume chamber 324, force isapplied to the piston and it begins to move to the right. Because piston322 is not yet in contact with mechanical clutch 340, piston 322 movesrelatively easily, and pressurized fluid entering clutch volume chamber324 achieves a relatively rapid movement of piston 322. During thisfirst phase, the volume of hydraulic fluid in the clutch volume chamber324 changes rapidly. A third phase can be defined once piston 322 movestoward the right extreme of travel and contacts mechanical clutch 340,force applied upon piston 322 transmits force to mechanical clutch 340and creates compressive pressure between clutch plate 345. Becausepiston 322 is subject to equal force from clutch plates 345 as thepiston is transmitting, piston 322 moves much more slowly as a result ofpressurized fluid acting upon the piston. In this third phase, becausethe piston only moves with additional compression of components ofmechanical clutch 340, the volume of hydraulic fluid in clutch volumechamber 324 changes more slowly. In the transitional period between thepiston in zero contact and the application of compressive force upon theclutch plates, a second phase can be defined wherein piston 322transitions from a period of relatively rapid movement and a period ofrelatively slower movement. Abrupt application of force upon clutchplates 345 can have adverse effects, including damage to the plates andpotential perception of the contact. Wave plate 328 can be used as partof mechanical clutch 340 to absorb some portion of force of the abruptcontact, making the transition between the first and third phases lessabrupt. Further, a touching state can be defined wherein between the endof the first phase and the initiation of the second phase, wherein thepiston begins to contact and exert force upon mechanical clutch 340 andwave plate 328.

As described above, clutches transition between locked and unlockedstates, and clutches designed to operate synchronously or without sliprequire substantially zero relative velocity when reactive torque istransmitted through the clutch. Strategies for synchronous operation ofclutches include synchronizing the clutch connective surfaces, thenapplying a clamping force to lock the clutch, thereby creating a clutchtorque capacity in the clutch, and then transmitting reactive torquethrough the clutch.

Clutch control strategies, sequentially, and in some instancesimultaneously, performing operations to synchronize the clutch plates,actuate the clutch to first the touching state and then to a fullylocked state, and then ramp up torque capacity of the locked clutch. Theorder in which these operations are performed are important tosynchronous operation, but also, the entire clutch transition must occurin as short a time span as possible to preserve drivability. Commandsmust be given to various powertrain components, accounting for reactiontimes, in order to generate the various operations involved in a shiftoccur in order with as little delay as possible. Commands and resultingactions can be simultaneous and overlapping, and understanding the timethat various components take to reach a particular state in response tocommands is important to coordinating the reactions in the orderrequired in synchronous clutch operation. Commands to the hydrauliccontrol system actuating the clutch and the resulting actions in theclutch are important to the sequential steps described above.

A pressure switch cycle time is a measure of the response times thatresult in clutch fill operations from the initiation of a hydraulic flowto a hydraulically actuated clutch until some clutch state of interest,as established by the configuration of the PCS controlling clutchoperation. Pressure switch cycle times, measuring a time to a PCSentering a full feed state to exiting the full feed state, wherein anoverlap state following the full feed state is intended to create atouching state in an associated clutch, can be utilized to diagnose atime until the clutch reaches a touching state through analysis of thetimes. An exemplary method to utilize pressure switch cycle times is totrack, first, a pressure switch signal, corresponding to a commandpressure initiating a full feed state and exceeding the calibratedpressure of the pressure switch, and, second, a pressure switch signal,corresponding to a drop in sensed pressure, for example, if the commandpressure is cut off from the pressure sensor through the plunger of thePCS reaching an overlap state. The time span between these two pressureswitch signals can be tracked as a pressure switch cycle time measuringthe time required to create an overlap state in the PCS. FIG. 6graphically illustrates fill times required to reach an overlap stateversus command pressure in a PCS, in accordance with the presentdisclosure.

The exemplary data of FIG. 6 demonstrates an overall trend in thepressure switch cycle times. Fill pressures resulting in clutch feedline and within a clutch volume chamber are the result of P_(LINE)applied to the PCS minus any pressure losses resulting from flow throughthe PCS and relates lines. Pressure losses resulting from flow can bedescribed by the following equation:

PRESSURE_LOSS=FLOW*FLOW_RESISTANCE   [3]

Flow resistance is a fixed term for a fixed geometry of the PCS at agiven setting. Pressure loss is therefore proportional to flow. Flowthrough the PCS to the clutch is high when the clutch piston is beingdisplaced, for example, in the first phase described above. Flow throughthe PCS to the clutch is low when the clutch piston is relativelystationary, for example, in the third phase described above, wherein thepiston is actively compressing the clutch plates. Applied to the trendin the data of FIG. 6, low command pressures correspond to low pressureswitch cycle times. Pressure switch cycle times relate the time betweenthe initiation and end of the full feed state. A low command pressure iseasily countered by the feedback pressure resulting in the PCS from theapplication of P_(LINE) in the full feed state. The resulting fillpressure in the clutch has only started to ramp up when the PCS is setto the overlap state, arresting any further increase in pressure withinthe clutch cylinder, so the PCS reaches the overlap state while theclutch is still in the first phase, described above. As the commandpressures increase, the pressure switch cycle times also quicklyincrease. Because the clutch is still in the first phase described abovewherein the piston is quickly displaced by hydraulic flow, the resultinghigh flow still results in high pressure losses. As a result, the fillpressure and the resulting pressure within the feedback loop riseslowly, resulting in significantly long incremental pressure switchcycle times with each incremental increase in the command pressure. Oncethe piston has been displaced and the clutch enters the third phase,described above, additional force applied to the piston throughapplication of fill pressure produces less movement, corresponding tocompression of the wave plate and clutch plates. Volume in the clutchvolume chamber changes slowly, and the resulting hydraulic flow into thecylinder is reduced. As a result, pressure losses resulting from flowdecrease, and the fill pressure rapidly approaches a static pressure orP_(LINE). Because the fill pressure in this third phase increasesrapidly, the time span needed to incrementally increase fill pressuredecreases. Increases in command pressure at higher command pressuresresult in only small increases in pressure switch cycle times.

As a result of the above behavior, low command pressures to the PCScorrespond to low pressure switch cycle times. As command pressuresincrease, the incremental times to reach an overlap state in the PCSincrease quickly at first, and then more slowly as the clutch platesbegin to compress. The transition between these two regions of pressureswitch cycle time behavior describes the transition between the firstphase and third phase, or the second phase. As described above, thetouching state occurs between the end of the first phase and theinitiation of the second phase.

As described above, command pressures in the steep section of the curvecorrespond to a PCS ceasing full feed state operation before thetouching state in the connected clutch is achieved. In a clutch shiftrequiring that a touching state be efficiently achieved, transitioningthe PCS to an overlap state before the touching state is reached is notpreferable. However, high fill pressures overshoot the touching state inthe clutch and can cause drivability issues. By analyzing a sample ofpressure switch cycle times through a range of command pressures,differences in the cycle times for incremental increases in the commandpressure can be used to calibrate or determine a preferred commandpressure to quickly and precisely produce a touching state.

FIG. 7 graphically illustrates fill times required to reach an overlapstate versus command pressure in a PCS through a range of incrementallydecreased command pressures, in accordance with the present disclosure.As described above in relation to FIG. 6, command pressures in the steepsection of the curve correspond to a PCS ceasing full feed stateoperation before the touching state in the connected clutch is achieved.Because this result in operation is not preferable and because overlyhigh command pressures can be anticipated and adjusted for, a method todetect a transition in incremental pressure switch cycle times beginningwith slightly high command pressures is preferred. Command pressures areshown as incrementally decreasing samples P₁ through P₆. Correspondingpressure switch cycle times are shown as T₁ through T₆. Changes betweenthe pressure switch cycle times are depicted as D₁ through D₅. Acomparison of D values yields little change between D₁ and D₃, withthese values belonging to a subset of D values with the shortest changesto pressure switch cycle times. However, D₄ is significantly increasedfrom D₃. This change indicates that a transition in pressure switchcycle times is indicated between P₄ and P₅. By selecting a preferredcommand pressure at or above P₄, the command pressure will not cease afull feed state prior to the touching state being achieved. Onepreferred method, to insure that the preferred command pressure isrobustly above the transition point indicated in FIG. 7, is to add acommand pressure adjustment to the first P value showing the increased Dvalue. An exemplary command pressure increase is one and a half timesthe value of the incremental decrease utilized in the calibrationsamples. In the present exemplary data, this would create a preferredcommand pressure half way between P₄ and P₃. In this way, pressureswitch cycle times can be compared through a range of command pressuresto calibrate or determine a preferred command pressure to efficientlyachieve a touching state.

The above methods to calibrate a preferred command pressure can beperformed once and maintained indefinitely for use in filling theclutch. However, with changing temperatures and wear in the system overtime, behavior of the clutch through a fill event can change.Characteristics of clutch fill events, such as measured pressure switchcycle times versus expected pressure switch cycle times, can be used tocontinually or periodically validate the preferred command pressure. Inthe event that the measure values differ from expected values by morethan a threshold, the calibration process can be reinitiated todetermine a new preferred command pressure. This process can occur anumber of times through the lifespan of a powertrain in order tomaintain an ability to precisely indicate a touching state in a clutchfill event.

FIG. 8 illustrates a flowchart describing an exemplary process to selectand continually validate preferred command pressures, in accordance withthe disclosure. Process 400 starts in step 402. At step 404, anuntrained state is initiated. At step 406, an old preferred commandpressure is purged, if it exists from a previous calibration. At step408, a training state is initiated. At step 410, an iterativelydecreasing command pressure calibration is performed, according tomethods described herein. At step 412, a preferred command pressureindicating occurrence of a touching state is selected, according tomethods described herein. At step 414, a trained state is initiated. Atstep 416, measured pressure switch cycle times are compared to expectedpressure switch cycle times. At step 418, the comparison of 416 isutilized to either validate or invalidate the preferred commandpressure. If the preferred command pressure remains validated, then theprocess reiterates to 416 wherein future measured fill times arecompared to expected fill times. If the preferred command pressure isinvalidated, then the process reiterates to step 404, wherein theselection of a new preferred command pressure begins.

The methods described herein can be performed in a PCS control modulelocated within a larger control system or located individually as aunitary device.

It is understood that modifications are allowable within the scope ofthe disclosure. The disclosure has been described with specificreference to the preferred embodiments and modifications thereto.Further modifications and alterations may occur to others upon readingand understanding the specification. It is intended to include all suchmodifications and alterations insofar as they come within the scope ofthe disclosure.

1. Method for controlling a powertrain comprising an electro-mechanicaltransmission mechanically-operatively coupled to an internal combustionengine and an electric machine adapted to selectively transmitmechanical power to an output member via selective application of aplurality of hydraulically-applied torque transfer clutches, said methodcomprising: applying through a series of clutch fill events a series ofincrementally changing command pressures in a pressure control solenoidcontrollably connected to one of said clutches; monitoring a pressureswitch fluidly connected to said pressure control solenoid andconfigured to indicate when said pressure switch is in a full feedstate; determining changes in cycle times of said pressure switchcorresponding to sequential applications of said series of incrementallychanging command pressures; selecting a preferred command pressure toachieve a transient state in said clutch based upon said changes inpressure switch cycle times; and controlling said clutch based upon saidpreferred command pressure.
 2. The method of claim 1, wherein applyingthrough said series of clutch fill events said series of incrementallychanging command pressures comprises: applying a series of incrementallydecreasing command pressures.
 3. The method of claim 1, whereinselecting said preferred command pressure based upon said changes inpressure switch cycle times comprises: selecting a lowest of saidcommand pressures conforming with a subset of shortest of said changesin cycle times of said pressure switch.
 4. The method of claim 1,wherein selecting said preferred command pressure based upon saidchanges in pressure switch cycle times comprises: identifying a first ofsaid command pressures to create an overlap state to correspond to achange in cycle times of said pressure switch not conforming with asubset of shortest of said changes in cycle times of said pressureswitch; and selecting as said preferred command pressure a valueequaling the sum of said identified first of said command pressures anda command pressure adjustment.
 5. The method of claim 4, wherein saidcommand pressure adjustment equals one and a half times said incrementalchange utilized in said series of incrementally changing commandpressures.
 6. The method of claim 1, further comprising: monitoring saidcontrolling said clutch based upon said preferred command pressure; andif said monitoring said controlling ceases to validate said preferredcommand pressure, selecting a new preferred command pressure.
 7. Themethod of claim 1, wherein selecting said preferred command pressurebased upon said changes in said pressure switch cycle times comprisesselecting said preferred command pressure based upon achieving atouching state in said clutch.
 8. Method for controlling a powertraincomprising an electro-mechanical transmission mechanically-operativelycoupled to an internal combustion engine and an electric machine adaptedto selectively transmit mechanical power to an output member viaselective application of a plurality of hydraulically-applied torquetransfer clutches, said method comprising: selecting a preferred commandpressure in a pressure control solenoid fluidly connected to one of saidclutches, said selecting comprising iteratively filling said clutch witha calibration command pressure, wherein said calibration commandpressure decreases with each iteration, monitoring pressure switch cycletimes for a pressure switch fluidly connected to said pressure controlsolenoid and configured to generate an indication when said pressurecontrol solenoid is in a full feed state, and selecting from saidcalibration command pressures said preferred command pressure based uponidentifying a transition between differences in said calibration commandpressures; and utilizing said preferred command pressure in subsequentoperation of said clutch.
 9. The method of claim 8, further comprising:comparing pressure switch cycles times during said subsequent operationof said clutch to expected cycle times; and reselecting said preferredcommand pressure based upon said comparing.
 10. An apparatus forcontrolling a powertrain comprising an electro-mechanical transmissionmechanically-operatively coupled to an internal combustion engine and anelectric machine adapted to selectively transmit mechanical power to anoutput member via selective application of a plurality ofhydraulically-applied torque transfer clutches, said method comprising:a pressure control solenoid fluidly connected to one of said clutches; apressure switch fluidly connected to said pressure control solenoidmonitoring when said pressure control solenoid is in a full feed state;a pressure control solenoid control module, including programming toapply through a series of clutch fill events a series of incrementallychanging command pressures in said pressure control solenoid to createan overlap state in said pressure control solenoid, monitor saidpressure switch, determine changes in cycle times of said pressureswitch corresponding to sequential applications of said series ofincrementally changing command pressures, select a preferred commandpressure based upon said changes in pressure switch cycle times, andcontrol said clutch based upon said preferred command pressure.
 10. Theapparatus of claim 9, wherein said series of incrementally changingcommand pressures comprises a series of incrementally decreasing commandpressures.
 11. The apparatus of claim 10, wherein said programming toselect said preferred command pressure based upon said changes inpressure switch cycle times comprises: selection of a lowest of saidcommand pressures conforming with a subset of shortest of said changesin cycle times of said pressure switch.
 12. The apparatus of claim 9,wherein said programming to select said preferred command pressure basedupon said changes in pressure switch cycle times is based upondetermining achieving a touching state in said clutch.
 13. The apparatusof claim 9, wherein said programming further comprises: validation ofsaid preferred command pressure through subsequent operation of saidclutch; and reselection of said preferred command pressure based upon afailure in said validation.